Internal Combustion Engine Control Device and Internal Combustion Engine Control System

ABSTRACT

In an internal combustion engine control device, when the engine stops, a variable valve actuation device is controlled to change an operating mode of each of intake valves to a specific state where an all-cylinder valve closed period, during which the intake valves of all engine cylinders are all kept in their non-lifted states, occurs, and a crank position change mechanism is controlled to change a crankshaft-rotation stopped position to a crankangle included within the all-cylinder valve closed period. When restarting the engine, the variable valve actuation device is controlled to bring the operating mode of each of the intake valves closer to an engine-start desired lift characteristic suited to an engine condition, prior to cranking action.

TECHNICAL FIELD

The present invention relates to a control device and a control system,capable of enhancing a restartability of a multi-cylinder internalcombustion engine.

BACKGROUND ART

In recent years, there have been proposed and developed variousmulti-cylinder engine control devices configured to enhance arestartability. One such multi-cylinder internal combustion enginecontrol device has been disclosed in Japanese Patent ProvisionalPublication No. 2003-56316 (hereinafter is referred to as“JP2003-056316”). Briefly speaking, as a variable valve actuationdevice, the engine control device disclosed in JP2003-056316 has acontinuous variable valve event and lift control (VEL) mechanism capableof continuously varying a valve lift amount and a working angle of anintake valve and a variable phase control mechanism (a variable valvetiming control (VTC) mechanism) capable of phase-shifting a centralphase angle (e.g., a phase at a peak valve lift) of a valve liftcharacteristic curve of the intake valve. When an intake-valve driveshaft is rotated in synchronism with rotation of an engine crankshaft,rotary motion of the intake-valve drive shaft is converted intooscillating motion of a rockable cam via the VEL mechanism. Alsoprovided is a hydraulically-operated valve-lash adjuster interleavedbetween the rockable cam and the intake-valve stem, so as to providezero valve lash, utilizing hydraulic pressure. In an engine stoppedstate, valve lift characteristics of intake valves of all enginecylinders are set to realize a so-called all-cylinder zero-lift statewhere an intake-valve lift amount of each individual engine cylinder, towhich the hydraulically-operated lash adjuster is applied, is almostzero.

That is to say, on the assumption that a stopped position of rotarymotion of a crankshaft of a four-cylinder internal combustion engine isa substantially midpoint between a piston top dead center (TDC) positionon compression stroke and a piston bottom dead center (BDC) position oncompression stroke, this crankshaft-rotation stopped position is set, sothat intake valves of all engine cylinders become kept at theirzero-lift states. Hence, in the engine stopped state, it is possible toprevent the hydraulically-operated lash adjuster of each individualcylinder from undesirably contracting owing to working oil leakage, andwhereby fluctuations in valve lift amounts between the engine cylinderscan be minimized, thus improving an engine restartability.

SUMMARY OF THE INVENTION

However, in the case of the engine control device as disclosed inJP2003-056316, as previously discussed, the intake-valve liftcharacteristic (in particular, intake-valve open timing and intake-valveclosure timing) of each individual engine cylinder is set to realize thepreviously-noted all-cylinder zero-lift state, on the assumption that acrankshaft-rotation stopped position is a substantially midpoint betweena TDC position on compression stroke and a BDC position on compressionstroke. Hence, in an engine stopped state, the peak-lift phase of theintake valve tends to become a phase near the TDC position (see FIG. 8of JP2003-056316) or a phase near the BDC position (see FIG. 9 ofJP2003-056316). Thus, when cranking the engine, the VTC mechanism mustbe operated greatly, but the VTC mechanism exhibits a bad operationalresponsiveness at very low engine speeds. Such a control device requiresa comparatively long time duration until an initial explosion of theengine takes place. As a result of this, it is difficult to ensure agood restartability.

It is, therefore, in view of the previously-described disadvantages ofthe prior art, an object of the invention to provide a control device ora control system configured to control a variable valve actuation devicein such a manner as to bring an actual lift characteristic of an intakevalve of each individual cylinder of an internal combustion enginecloser to a desired lift characteristic suited to a monitored enginecondition, prior to cranking action, thereby ensuring a good enginerestartability.

In order to accomplish the aforementioned and other objects of thepresent invention, an internal combustion engine control devicecomprises a crank position change mechanism configured to change acrankshaft-rotation stopped position of a crankshaft of an internalcombustion engine, and a variable valve actuation device configured tochange at least a working angle of each of intake valves of a pluralityof engine cylinders by changing a position of a control shaft, wherein,when the engine stops, the variable valve actuation device is controlledto change an operating mode of each of the intake valves to a specificstate where an all-cylinder valve closed period, during which the intakevalves of the cylinders are all kept in their non-lifted states, occurs,and the crank position change mechanism is controlled to change thecrankshaft-rotation stopped position to a crankangle included within theall-cylinder valve closed period, and wherein, when restarting theengine, the variable valve actuation device is controlled to bring theoperating mode of each of the intake valves closer to an engine-startdesired lift characteristic suited to an engine condition, prior tocranking action.

According to another aspect of the invention, an internal combustionengine control device comprises a crank position change mechanismconfigured to change a crankshaft-rotation stopped position of acrankshaft of an internal combustion engine, and a variable valveactuation device configured to change at least a working angle of eachof intake valves of a plurality of engine cylinders by changing aposition of a control shaft, wherein, when the engine stops, the crankposition change mechanism, together with the variable valve actuationdevice, is controlled to execute crank position control as well asintake-valve operating characteristic control in such a manner as torealize a specific state where there is a less valve-spring reactionforce acting on the control shaft, and wherein, when restarting theengine, the variable valve actuation device is controlled to bring theposition of the control shaft closer to a desired position suited tostart-up of the engine, prior to cranking action.

According to a further aspect of the invention, an internal combustionengine control system comprises a crank position change mechanismconfigured to change a crankshaft-rotation stopped position of acrankshaft of an internal combustion engine, a variable valve actuationdevice configured to change a valve lift as well as a working angle ofeach of intake valves of a plurality of engine cylinders, and acontroller configured to control operations of the crank position changemechanism and the variable valve actuation device, wherein, when theengine stops, the controller controls the variable valve actuationdevice to change an operating characteristic of each of the intakevalves to a specific state where an all-cylinder valve closed period,during which the intake valves of the cylinders are all kept in theirnon-lifted states, occurs, and the controller controls the crankposition change mechanism to change the crankshaft-rotation stoppedposition to a crankangle included within the all-cylinder valve closedperiod, and wherein, when restarting the engine, the controller controlsthe variable valve actuation device to bring an engine-start desiredlift characteristic of each of the intake valves, suited to an enginecondition, prior to cranking action.

The other objects and features of this invention will become understoodfrom the following description with reference to the accompanyingdrawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a schematic system diagram illustrating an internal combustionengine to which a variable valve actuation device of an embodiment canbe applied.

FIG. 2 is a perspective view illustrating the variable intake-valveactuation device of the embodiment, which includes a continuouslyvariable valve event and lift control (VEL) mechanism and a variablevalve timing control (VTC) mechanism.

FIGS. 3A-3B are axial views showing the operation of the intake-valveVEL mechanism during a small-lift control mode.

FIGS. 4A-4B are axial views showing the operation of the intake-valveVEL mechanism during a large-lift control mode.

FIG. 5 is a variable intake-valve lift and event (working angle) andphase characteristic diagram, obtained by both of the intake-valve VELand VTC mechanisms of the variable valve actuation device of theembodiment.

FIG. 6 is a cross-sectional view showing the intake-valve VTC mechanismincluded in the variable valve actuation device of the embodiment.

FIG. 7 is a lateral cross-section taken along the line A-A of FIG. 6,and showing the maximum phase-retard state of the intake-valve VTCmechanism.

FIG. 8 is a lateral cross-section taken along the line A-A of FIG. 6,and showing the maximum phase-advance state of the intake-valve VTCmechanism.

FIGS. 9A-9D are characteristic diagrams showing the relationship amongcrankangle, intake-valve open timing, and intake-valve closure timing,obtained by the variable valve actuation device of the first embodiment,at each of #1, #3, #4, and #2 engine cylinders.

FIG. 10 is a flowchart showing a control routine executed within acontroller incorporated in the engine control system of the presentembodiment.

FIGS. 11A-11B are characteristic diagrams showing the relationship amongcrankangle, intake-valve open timing, and intake-valve closure timing,obtained by the variable valve actuation device of the secondembodiment, at each of #1, and #2 engine cylinders.

FIG. 12 is an axial view showing the operation of the intake-valve VELmechanism included in the variable valve actuation device of the secondembodiment.

DESCRIPTION OF THE PREFERRED EMBODIMENTS First Embodiment

Referring now to the drawings, particularly to FIGS. 1-2, the enginecontrol device of the first embodiment is exemplified in a four-cyclefour-cylinder internal combustion gasoline engine having four valves percylinder, namely two intake valves 4, 4 (see FIGS. 1-2) and two exhaustvalves 08, 08 (see FIG. 1).

The construction of the multiple-cylinder internal combustion engine, towhich the engine control device of the embodiment can be applied, ishereunder described in detail in reference to the system diagram ofFIG. 1. Also, the engine control device of the embodiment can be appliedto a hybrid vehicle (HV) employing an automatic engine stop-restartsystem capable of temporarily automatically stopping an internalcombustion engine during idling without depending on a driver's will, inaddition to an internal-combustion-engine equipped vehicle with aso-called idling-stop system that enables an idling-stop action. Theengine of FIG. 1 is constructed by a cylinder block SB having a cylinderbore, a reciprocating piston 01 movable or slidable through a stroke inthe cylinder bore, a cylinder head SH on the cylinder block SB, anintake port IP and an exhaust port EP formed in cylinder head SH, twointake valves 4, 4 each slidably installed on cylinder head SH foropening and closing the opening end of intake port IP, and two exhaustvalves 08, 08 each slidably installed on cylinder head SH for openingand closing the opening end of exhaust port EP.

Piston 01 is connected to an engine crankshaft 02 via a connecting rod03. A combustion chamber 04 is defined between the piston crown ofpiston 01 and the underside of cylinder head SH.

An electronically-controlled throttle valve unit SV is provided upstreamof intake port IP and located in an interior space of an intake manifoldIa of an intake pipe I connected to intake port IP, for controlling aquantity of intake air. A fuel injector or a fuel injecting valve (notshown) is provided downstream of throttle valve unit SV. A spark plug 05is located substantially in a middle of cylinder head SH.

A flywheel ring gear 09 of a flywheel attached to engine crankshaft 02is in meshed-engagement with a pinion gear mechanism 06. Cranking actionof crankshaft 02 can be initiated by driving the pinion gear mechanism06 by means of an electric motor (or a cranking motor) 07, andsimultaneously a rotational position of crankshaft 02 can be controlled.That is, electric motor 07 and pinion gear mechanism 06 construct a partof a crank position change mechanism.

Each of intake valves 4, 4 is forced by the associated valve spring 5 ina direction that the opening end of intake port IP is closed.

As clearly shown in FIGS. 1-2, particularly, in FIG. 2, the variablevalve actuation device incorporated in the engine control system of theembodiment is comprised of an intake-valve variable valve event and liftcontrol (VEL) mechanism 1 and an intake-valve variable valve timingcontrol (VTC) mechanism (or a variable phase control mechanism) 2.Intake-valve VEL mechanism 1 is able to simultaneously control or adjustor change both of a valve lift and a lifted-period (a working angle, inother words, a valve open period) of each of intake valves 4, 4.Intake-valve VTC mechanism 2 is able to advance or retard only a phaseof each of intake valves 4, 4, while keeping a valve lift and workingangle characteristic of each intake valve 4 constant. In the shownembodiment, there is no exhaust-valve VTC mechanism in the exhaust valveside, and thus exhaust-valve open timing (EVO) and exhaust valve closuretiming (EVC) are both fixed.

As the VEL mechanism 1, the variable valve actuation system of theembodiment uses a continuously variable valve event and lift controlmechanism as disclosed in Japanese Patent Provisional Publication No.2003-172112. Briefly speaking, as shown in FIG. 2, VEL mechanism 1 iscomprised of a cylindrical hollow drive shaft 6, a ring-shaped drive cam7, two rockable cams 9, 9, and a multinodular-link motion transmittingmechanism (or a motion converter) mechanically linked between drive cam7 and the rockable-cam pair (9, 9) for transmitting a torque created bydrive cam (eccentric cam) 7 as an oscillating force of each of rockablecams 9, 9. Cylindrical hollow drive shaft 6 is rotatably supported bybearings in the upper part of cylinder head SH. Drive cam 7 is formed asan eccentric cam that is press-fitted or integrally connected onto theouter periphery of drive shaft 6. Rockable cams 9, 9 are oscillatinglyor rockably supported on the outer periphery of drive shaft 6 and insliding-contact with respective upper contact surfaces of two valvelifters 8, 8, which are located at the valve stem ends of intake valves4, 4. In other words, the motion transmitting mechanism (or the motionconverter) is provided to convert a rotary motion (input torque) ofdrive cam 7 into an up-and-down motion (a valve opening force) of eachintake valve 4 (i.e., an oscillating force creating an oscillatingmotion of each rockable cam 9).

Torque is transmitted from engine crankshaft 02 through a timingsprocket 33 fixedly connected to one axial end of drive shaft 6 via atiming chain (not shown) to the drive shaft 6. As indicated by the arrowin FIG. 2, the direction of rotation of drive shaft 6 is set in aclockwise direction.

Drive cam 7 has an axial bore that is displaced from the geometriccenter of the cylindrical drive cam 7. Drive cam 7 is fixedly connectedto the outer periphery of drive shaft 6, so that the inner peripheralsurface of the axial bore of drive cam 7 is press-fitted onto the outerperiphery of drive shaft 6. Thus, the center of drive cam 7 is offsetfrom the shaft center of drive shaft 6 in the radial direction by apredetermined eccentricity (or a predetermined offset distance).

As best seen from the axial views shown in FIGS. 2, 3A-3B and 4A-4B,each of rockable cams 9, 9 is formed as a substantially raindrop-shapedcam. Rockable cams 9, 9 have the same cam profile. Rockable cams 9, 9are formed integral with respective axial ends of a cylindrical-hollowcamshaft 10. Cylindrical-hollow camshaft 10 is rotatably supported ondrive shaft 6. The outer peripheral contacting surface of rockable cam9, in sliding-contact with the upper contact surface of valve lifter 8,includes a cam surface 9 a. The base-circle portion of rockable cam 9 isintegrally formed with or integrally connected to camshaft 10, to permitoscillating motion of rockable cam 9 on the axis of drive shaft 6. Theouter peripheral surface (cam surface 9 a) of rockable cam 9 isconstructed by a base-circle surface, a circular-arc shaped ramp surfaceextending from the base-circle surface to a cam-nose portion, atop-circle surface (simply, a top surface) that provides a maximum valvelift (or a maximum lift amount), and a lift surface by which the rampsurface and the top surface are joined. The base-circle surface, theramp surface, the lift surface, and the top surface abut predeterminedpositions of the upper surface of valve lifter 8, depending on theoscillatory position of rockable cam 9.

The motion transmitting mechanism (the motion converter) is comprised ofa rocker arm 11 laid out above drive shaft 6, a link arm 12 mechanicallylinking one end (or a first end portion 11 a) of rocker arm 11 to thedrive cam 7, and a link rod 13 mechanically linking the other end (asecond end portion 11 b) of rocker arm 11 to the cam-nose portion ofrockable cam 9.

Rocker arm 11 is formed with an axially-extending center bore (a throughopening). The rocker-arm center bore of rocker arm 11 is rotatablyfitted onto the outer periphery of a control cam 18 (described later),to cause a pivotal motion (or an oscillating motion) of rocker arm 11 onthe axis of control cam 18. The first arm portion 11 a of rocker arm 11extends from the axial center bore portion in a first radial direction,whereas the second arm portion lib of rocker arm 11 extends from theaxial center bore portion in a second radial direction substantiallyopposite to the first radial direction. The first arm portion 11 a ofrocker arm 11 is rotatably pin-connected to link arm 12 by means of aconnecting pin 14, while the second arm portion 11 b of rocker arm 11 isrotatably pin-connected to one end (a first end 13 a) of link rod 13 bymeans of a connecting pin 15.

Link arm 12 is comprised of a comparatively large-diameter annular baseportion 12 a and a comparatively small-diameter protruding end portion12 b radially outwardly extending from a predetermined portion of theouter periphery of large-diameter annular base portion 12 a.Large-diameter annular base portion 12 a is formed with a drive-camretaining bore, which is rotatably fitted onto the outer periphery ofdrive cam 7. On the other hand, small-diameter protruding end portion 12b of link arm 12 is pin-connected to the first arm portion 11 a ofrocker arm 11 by means of connecting pin 14.

Link rod 13 is pin-connected at the other end (a second end 13 b) to thecam-nose portion of rockable cam 9 by means of a connecting pin 16.

Also provided is a motion-converter attitude control mechanism thatchanges an initial actuated position (a fulcrum of oscillating motion ofrocker arm 11) of the motion transmitting mechanism (or the motionconverter). As clearly shown in FIGS. 3A-3B and 4A-4B, the attitudecontrol mechanism includes a control shaft 17 and control cam 18.Control shaft 17 is located above and arranged in parallel with driveshaft 6 in such a manner as to extend in the longitudinal direction ofthe engine, and rotatably supported on cylinder head SH by means of thesame bearing members as drive shaft 6. Control cam 18 is attached to theouter periphery of control shaft 17 and slidably fitted into andoscillatingly supported in a control-cam retaining bore formed in rockerarm 11. Control cam 18 serves as a fulcrum of oscillating motion ofrocker arm 11. Control cam 18 is integrally formed with control shaft17, so that control cam 18 is fixed onto the outer periphery of controlshaft 17. Control cam 18 is formed as an eccentric cam having acylindrical cam profile. The axis (the geometric center) of control cam18 is displaced a predetermined distance from the axis of control shaft17.

As shown in FIG. 2, the attitude control mechanism also includes a drivemechanism 19. Drive mechanism 19 is comprised of a geared motor or anelectric control-shaft actuator 20 fixed to one end of a housing (notshown) and a ball-screw motion-transmitting mechanism (simply, aball-screw mechanism) 21 that transmits a motor torque created byelectric motor (simply, motor) 20 to control shaft 17. In more detail,motor 20 is constructed by a proportional control type direct-current(DC) motor. Motor 20 is controlled in response to a control signal,which is generated from the output interface circuitry of an electroniccontrol unit (simply, a controller) 22 and whose signal value isdetermined based on engine/vehicle operating conditions.

Ball-screw mechanism 21 is comprised of a ball-screw shaft (or a wormshaft) 23 coaxially aligned with and connected to the motor output shaftof motor 20, a substantially cylindrical, movable ball nut 24 threadablyengaged with the outer periphery of ball-screw shaft 23, a link arm 25fixedly connected to the rear end 17 a of control shaft 17, a linkmember 26 mechanically linking link arm 25 to ball nut 24, andrecirculating balls interposed between the worm teeth of ball-screwshaft 23 and guide grooves cut in the inner peripheral wall surface ofball nut 24. In a conventional manner, a rotary motion (input torque) ofball-screw shaft 23 is converted into a rectilinear motion of ball nut24 through the recirculating balls. Ball nut 24 is axially forced towardmotor 20 by the spring force of a return spring (a coil spring) 30,serving as a biasing device (biasing means). The direction of the springforce (spring bias) of return spring 30 corresponds to a direction thatthe VEL mechanism is biased to a minimum valve lift and working anglecharacteristic. Hence, during an engine stopping period, ball nut 24 ofball-screw mechanism 21 can be forced and displaced by the spring biasof coil spring 30 in the axial direction of ball-screw shaft 23corresponding to the minimum valve lift and working angle characteristicof the intake-valve VEL mechanism.

As shown in FIGS. 1-2, controller 22 generally comprises amicrocomputer. Controller 22 includes an input/output interface (I/O),memories (RAM, ROM), and a microprocessor or a central processing unit(CPU). The input/output interface (I/O) of controller 22 receives inputinformation from various engine/vehicle switches and sensors, namely acrankangle sensor (or a crankshaft position sensor) 27, an engine speedsensor, an accelerator opening sensor, a vehicle speed sensor, a rangegear position switch, a drive-shaft angular position sensor 28, acontrol-shaft angular position sensor 29, an airflow meter AFM, anengine temperature sensor (i.e., an engine coolant temperature sensor,and the like. Within ECU 22, the central processing unit (CPU) allowsthe access by the I/O interface of input informational data signals fromthe previously-discussed engine/vehicle switches and sensors. Theprocessor of ECU 22 determines the current engine/vehicle operatingcondition, based on input information from the engine/vehicle switchesand sensors. Crankangle sensor 27 is provided to detect an angularposition (crankangle) of crankshaft 02. Drive-shaft angular positionsensor 28 is provided for detecting an angular position of drive shaft6. Also, based on both of the sensor signals from crankangle sensor 27and drive-shaft angular position sensor 28, an angular phase of driveshaft 6 relative to timing sprocket 33 is detected. Control-shaftangular position sensor 29 is provided to detect an angular position ofcontrol shaft 17. Airflow meter AFM is provided for measuring ordetecting a quantity of air flowing through intake pipe I, andconsequently for detecting or estimating the magnitude of engine load.The CPU of ECU 22 is responsible for carrying the control program storedin memories and is capable of performing necessary arithmetic and logicoperations, for example, electronic throttle opening control achievedthrough the throttle actuator of electronically-controlled throttlevalve unit SV, electronic fuel injection control achieved by theelectronic fuel-injection system, electronic spark control achieved bythe electronic ignition system, valve lift and working angle controlexecuted by VEL mechanism 1, and phase control executed by VTC mechanism2. Computational results (arithmetic calculation results), that is,calculated output signals are relayed through the output interfacecircuitry of ECU 22 to output stages, namely the throttle actuator ofelectronically-controlled throttle valve unit SV,electronically-controlled fuel injectors of the fuel-injection system,electronically-controlled spark plugs 05 of the electric ignitionsystem, motor 20 of VEL mechanism 1, the solenoid of a directionalcontrol valve 47 (described later) for VTC mechanism 2, and electricmotor 07 used for cranking motor control.

Hereunder described briefly in reference to FIGS. 2, 3A-3B, 4A-4B, and 5is the basic operation of intake-valve VEL mechanism 1. In an engineoperating range such as in a low-speed low-load range, the control-shaftactuator (motor 20) of VEL mechanism 1 is driven in one rotationaldirection in response to a control signal generated from the outputinterface circuitry of ECU 22. Thus, ball-screw shaft 23 is rotated in adirection corresponding to the one rotational direction of motor 20 byinput torque created by motor 20, thereby producing a maximumrectilinear motion of ball nut 24 in one ball-nut axial direction thatball nut 24 approaches close to motor 20. As a result, control shaft 17rotates in one rotational direction via a linkage comprised of linkmember 26 and link arm 25.

As can be seen from the angular position of control cam 18 shown inFIGS. 3A-3B, by way of revolving motion of the center of control cam 18around the center of control shaft 17, the radially thick-walled portionof control cam 18 shifts upwards apart from drive shaft 6 and is held atthe upwardly shifted position, with the result that the pivot (theconnected point by connecting pin 15) between the second arm portion 11b of rocker arm 11 and the first rod end 13 a of link rod 13 also shiftsupwards with respect to drive shaft 6. As a result, the cam-nose portionof each of rockable cams 9, 9 is forcibly pulled up via the second rodend 13 b of link rod 13. As viewed from the front end of drive shaft 6,the angular position of each rockable cam 9 shown in FIGS. 3A-3B isrelatively shifted to the clockwise direction from the angular positionof each rockable cam 9 shown in FIGS. 4A-4B.

With control cam 18 held at the angular position shown in FIGS. 3A-3B,when drive cam 7 is rotated, a rotary motion of drive cam 7 is convertedthrough link arm 12, the first arm portion 11 a of rocker arm 11, thesecond arm portion 11 b of rocker arm 11, and link rod 13 into anoscillating motion of rockable cam 9, but almost the base-circle surfacearea of rockable cam 9 is brought into sliding-contact with the uppercontact surface of valve lifter 8 (see FIGS. 3A-3B). Thus, the actualintake-valve lift becomes a small lift L1 and simultaneously the actualintake-valve working angle becomes a small working angle D1 (see thesmall intake-valve lift L1 and small working angle D1 characteristicshown in FIG. 5).

Actually, a valve clearance exists between rockable cam 9 and valvelifter 8. Thus, a valve lift amount of intake valve 4 becomes less thana cam lift amount by the valve clearance, and therefore a lifted periodfrom valve open timing to valve closure timing, fully taking intoaccount the valve clearance, can be regarded as a working angle.

When shifting from the previously-noted low-speed low-load range toanother engine operating condition such as mid-speed and mid-loadoperation, motor 20 is rotated in a reverse-rotational directionresponsively to a control signal, which is generated from the outputinterface circuitry of ECU 22. Thus, ball-screw shaft 23 is also rotatedin the reverse-rotational direction by reverse-rotation of the motoroutput shaft of motor 20, thereby producing the opposite rectilinearmotion of ball nut 24 against the bias of coil spring 30. As a result,control shaft 17 rotates in the opposite rotation direction via thelinkage (25, 26).

By way of revolving motion of the center of control cam 18 around thecenter of control shaft 17, the radially thick-walled portion of controlcam 18 slightly downwardly shifts toward drive shaft 6 and is held atthe slightly downwardly shifted position. Thus, the attitude of rockerarm 11 slightly shifts anticlockwise from the angular position of rockerarm 11 shown in FIGS. 3A-3B, with the result that the pivot (theconnected point by connecting pin 15) between the second arm portion libof rocker arm 11 and the first rod end 13 a of link rod 13 also shiftsslightly downwards. As a result, the cam-nose portion of each ofrockable cams 9, 9 is forcibly slightly pushed down via the second rodend 13 b of link rod 13. As viewed from the front end of drive shaft 6,the angular position of each rockable cam 9 is relatively shifted to theanticlockwise direction from the angular position of each rockable cam 9shown in FIGS. 3A-3B.

With control cam 18 shifted from the angular position shown in FIGS.3A-3B to the intermediate angular position located in a substantiallymiddle of the angular position shown in FIGS. 3A-3B and the angularposition shown in FIGS. 4A-4B, during rotation of drive cam 7, a rotarymotion of drive cam 7 is converted through link arm 12, the first armportion 11 a of rocker arm 11, the second arm portion 11 b of rocker arm11, and link rod 13 into an oscillating motion of rockable cam 9. Atthis time, a part of the base-circle surface area, the ramp surfacearea, the lift surface area, and the top surface area are brought intosliding-contact with the upper contact surface of valve lifter 8. Thus,when varying from the angular position of control cam 18 shown in FIGS.3A-3B to the intermediate angular position, the actual intake-valve liftand working angle characteristic can be quickly varied from the smallintake-valve lift L1 and small working angle D1 characteristic to amiddle intake-valve lift L2 and middle working angle D2 characteristic(see FIG. 5). That is, intake-valve working angle as well asintake-valve lift can be simultaneously increased. Owing to a valve liftincrease (L1→L2) and a working angle increase (D1→D2), intake valveclosure timing IVC of each of intake valves 4, 4 is phase-retarded andcontrolled to a timing value near BDC. Thus, an effective compressionratio becomes high to ensure good combustion. Additionally, a chargingefficiency of fresh air tends to become high, thus resulting in anincrease in torque generated by combustion.

After this, when the engine/vehicle operating condition is shifting fromthe low or middle load range to a high-speed high-load range, motor 20is further driven in the reverse-rotational direction responsively to acontrol signal, which is generated from the output interface circuitryof ECU 22 and determined based on the high engine load condition. Thus,ball-screw shaft 23 is further rotated in the reverse-rotationaldirection by reverse-rotation of the motor output shaft of motor 20,thereby producing the further opposite rectilinear motion of ball nut24. As a result, control shaft 17 further rotates in the oppositerotation direction via the linkage (25, 26). By way of further revolvingmotion of the center of control cam 18 around the center of controlshaft 17, the radially thick-walled portion of control cam 18 furthershifts downwards and is held at the downwardly shifted position. Thus,the attitude of rocker arm 11 further shifts anticlockwise, with theresult that the pivot (the connected point by connecting pin 15) betweenthe second arm portion 11 b of rocker arm 11 and the first rod end 13 aof link rod 13 further shifts downwards. As a result, the cam-noseportion of each of rockable cams 9, 9 is further forcibly pushed downvia the second rod end 13 b of link rod 13. As viewed from the front endof drive shaft 6, the angular position of each rockable cam 9 is furthershifted anticlockwise. With control cam 18 shifted to the angularposition (suited to high load operation) shown in FIGS. 4A-4B, duringrotation of drive cam 7, a rotary motion of drive cam 7 is convertedthrough the motion transmitting mechanism (links 11, 12, and 13) into anoscillating motion of rockable cam 9. At this time, a part of thebase-circle surface area, the ramp surface area, the lift surface area,and the top surface area are brought into sliding-contact with the uppercontact surface of valve lifter 8. Thus, when switching from theintermediate angular position (suited to middle load operation) ofcontrol cam 18 to the angular position (suited to high load operation)shown in FIGS. 4A-4B, the actual intake-valve lift and working anglecharacteristic can be continuously varied from the middle intake-valvelift L2 and middle working angle D2 characteristic to a largeintake-valve lift L3 and large working angle D3 characteristic (see FIG.5). As a result, in a high-speed high-load range, a charging efficiencyof fresh air tends to become higher, thus more greatly enhancing enginepower output.

As can be appreciated from a plurality of intake-valve lift L andintake-valve working angle D characteristic curves (or a plurality ofintake-valve lift L and lifted-period D characteristic curves) shown inFIG. 5, according to VEL mechanism 1 incorporated in the internalcombustion engine control system of the embodiment, through all engineoperating conditions from low-speed low-load operation to high-speedhigh-load operation, the intake-valve lift and working anglecharacteristic can be continuously controlled or adjusted from the smallintake-valve lift L1 and working angle D1 characteristic via the middleintake-valve lift L2 and working angle D2 characteristic to the largeintake-valve lift L3 and working angle D3 characteristic, or vice versa.That is to say, the intake-valve lift and working angle characteristiccan be controlled or adjusted to an optimal characteristic suited to thelatest up-to-date information concerning engine operating condition.

During an engine stopping period, as previously described, ball nut 24of ball-screw mechanism 21 is forced and stably held by the spring biasof coil spring 30 in the axial direction of ball-screw shaft 23corresponding to the small lift L1 and small working angle D1 range.Hence, it is possible to reduce a frictional loss in the valve operatingsystem, thus ensuring a good startability (or a good restartability).

As discussed above, as best seen from the variable intake-valve lift andevent (working angle) and phase characteristic diagram of FIG. 5, aslight valve clearance A exists between the base-circle surface ofrockable cam 9 oscillating and the valve lifter 8. Hence, the effectivevalve lift amount L tends to be reduced from the cam lift amount of therockable cam by the valve clearance A, and thus the effective workingangle D also tends to be reduced slightly. The previously-discussedvalve lift amounts L1 to L3 and working angles D1 to D3 represent theeffective valve lifts and working angles, from which the valve clearanceA is excepted.

Hereunder described briefly in reference to FIGS. 6, 7, and 8, is theconstruction of intake-valve VTC mechanism 2. As can be appreciated fromthe cross sections of FIGS. 6-8, VTC mechanism 2 comprises a so-calledhydraulically-operated rotary vane type VTC mechanism. Intake-valve VTCmechanism 2 is comprised of timing sprocket 33 driven by crankshaft 02and fixedly connected to drive shaft 6 for torque transmission, afour-blade vane member 32 fixedly connected or bolted to the shaft endof drive shaft 6 and rotatably accommodated in the internal space oftiming sprocket 33, and a hydraulic circuit, which hydraulicallyoperates vane member 32 in a manner so as to rotate vane member 32 inselected one of normal-rotational and reverse-rotational directions.

Timing sprocket 33 is comprised of a substantially cylindrical,phase-converter housing 34 rotatably accommodating therein vane member32, a disk-shaped front cover 35 hermetically covering the front openingend of housing 34, and a disk-shaped rear cover 36 hermetically coveringthe rear opening end of housing 34. Housing 34 and front and rear covers35-36 are axially connected integral with each other by tightening fourbolts 37.

Housing 34 is substantially cylindrical in shape and opened at bothaxial ends. Housing 34 has four shoes 34 a, 34 a, 34 a, 34 a evenlyspaced around its entire circumference and serving as four partitionwalls radially inwardly extending from the inner periphery of thehousing.

Each of shoes 34 a is trapezoidal in shape, and has an axially-extendingbolt insertion hole 34 b formed in its substantially central portionsuch that bolt 37 is inserted into the bolt insertion hole. As best seenin FIG. 7, each of shoes 34 a has an axially-elongated seal grooveformed in its apex. Four elongated oil seals 38, 38, 38, 38 each havinga substantially C-shape in lateral cross section, are fitted into andretained in the respective seal grooves of shoes 34 a. Although it isnot clearly shown in FIG. 7, actually, four leaf springs are fitted intoand retained in the respective seal grooves of shoes 34 a in such amanner as to radially inwardly force the respective oil seals 38 againstthe outer peripheral wall surface of a vane rotor 32 a (describedlater).

The previously-noted disk-shaped front cover 35 has a comparativelylarge-diameter center supporting bore 35 a and circumferentiallyequidistant-spaced bolt holes (not numbered) bored to axially conform tothe respective bolt insertion holes 34 b of shoes 34 a of housing 34.

The previously-noted disk-shaped rear cover 36 is integrally formed atits rear end with a toothed portion 36 a, which is in meshed-engagementwith the timing chain. Also, rear cover 36 has a substantially centerbearing bore 36 b having a comparatively large diameter.

Vane member 32 is comprised of a substantially annular ring-shaped vanerotor 32 a formed with a center bolt insertion hole andradially-extending four vanes or blades 32 b, 32 b, 32 b, 32 b evenlyspaced around the entire circumference of vane rotor 32 a and integrallyformed on the outer periphery of vane rotor 32 a.

A small-diameter, cylindrical-hollow front end portion of vane rotor 32a is rotatably supported in the center bore 35 a of front cover 35. Asmall-diameter, cylindrical-hollow rear end portion of vane rotor 32 ais also rotatably supported in the bearing bore 36 b of rear cover 36.

Vane rotor 32 a of vane member 32 has an axially-extending central bore14 a into which a vane mounting bolt 39 b is inserted for bolting vanemember 32 to the front axial end of drive shaft 6 by axially tighteningvane mounting bolt 39 b.

One of four vane blades 32 b, 32 b, 32 b, 32 b is configured to have aninverted trapezoidal shape in lateral cross section, whereas theremaining three vane blades are configured to be substantiallyrectangular in lateral cross section. The remaining three blades havealmost the same circumferential width and the same radial length. Thecircumferential width of the one blade having the inverted trapezoidalshape is dimensioned to be greater than that of each of the remainingthree rectangular blades, taking account of total weight balance of vanemember 32, in other words, reduced rotational unbalance of vane member32 having four blades 32 b.

Each of four blades 32 b, 32 b, 32 b, 32 b is disposed in an internalspace defined between the associated two adjacent shoes 34 a and 34 a.As best seen in FIG. 7, four apex seals 40, 40, 40, and 40, each beingsubstantially C-shaped in lateral cross section, are fitted into andretained in respective seal grooves formed in apexes of four blades 32b, so that each of blades 32 b is slidable along the inner peripheralwall surface of phase-converter housing 34. Although it is not clearlyshown in FIG. 7, actually, four leaf springs are fitted into andretained in the respective seal grooves of the apexes of blades 32 b insuch a manner as to radially inwardly force the respective apex seals 40against the inner peripheral wall surface of housing 34. The backwardsidewall surface of each blade 32 b, opposing to the rotationaldirection of drive shaft 6, is formed with substantially circular, twoconcave grooves 32 c and 32 c, which serve as spring retaining holes fortwo rows of return springs 55-56. Return springs 55-56 are disposedbetween the spring-retaining-hole equipped backward sidewall surface ofblade 32 b and a spring-retaining sidewall surface of shoe 34 a opposingto the backward sidewall surface of blade 32 b.

Four blades 32 b of vane member 32 and four shoes 34 a of housing 34cooperate with each other to define four variable-volume phase-advancechambers 41 and four variable-volume phase-retard chambers 42. In moredetail, each of phase-advance chambers 41 is defined between thespring-retaining-hole equipped backward sidewall surface of blade 32 band the opposing spring-retaining sidewall surface of shoe 34 a. Each ofphase-retard chambers 42 is defined between thenon-spring-retaining-hole equipped forward sidewall surface of blade 32b and the opposing non-spring-retaining sidewall surface of shoe 34 a.

As clearly shown in FIG. 6, the previously-noted hydraulic circuit iscomprised of a first hydraulic line 43 provided to supply and exhaustworking fluid (hydraulic pressure) to and from each of phase-advancechambers 41, and a second hydraulic line 44 provided to supply andexhaust working fluid (hydraulic pressure) to and from each ofphase-retard chambers 42. That is, the hydraulic circuit comprises adual hydraulic line system (43, 44). Each of hydraulic lines 43 and 44are connected through an electromagnetic solenoid-operated directionalcontrol valve 47 to a working-fluid supply passage 45 and aworking-fluid drain passage 46. A one-way oil pump 49 is disposed insupply passage 45 for sucking working fluid in an oil pan 48 and fordischarging the pressurized working fluid from its discharge port. Thedownstream end of drain passage 46 communicates oil pan 48.

First and second hydraulic lines 43 and 44 are formed in a substantiallycylindrical flow-passage structure 39. One end (i.e., a first end) offlow-passage structure 39 is inserted through the left-hand axialopening end of the small-diameter, cylindrical-hollow front end portionof vane rotor 32 a into a cylindrical bore 32 d formed in vane rotor 32a. The other end (i.e., a second end) of flow-passage structure 39 isconnected to electromagnetic solenoid-operated directional control valve47. Three annular seals 39 s, 39 s, 39 s are disposed between the outerperiphery of the first end of flow-passage structure 39 and the innerperiphery of cylindrical bore 32 d of vane rotor 32 a. In more detail,annular seals 39 s are fitted into and retained in respective sealgrooves formed in the outer periphery of the first end of flow-passagestructure 39. These annular seals 39 s act to partition between aphase-advance-chamber communication port of first hydraulic line 43 anda phase-retard-chamber communication port of second hydraulic line 44 ina fluid-tight fashion.

First hydraulic line 43 is further provided with a working-fluid chamber43 a and four branch passages 43 b, 43 b, 43 b, 43 b. First hydraulicline 43 penetrates through the first end face of flow-passage structure39, and the axial passage of first hydraulic line 43 communicatesworking-fluid chamber 43 a. Working-fluid chamber 43 a is formed as theinner half of cylindrical bore 32 d of vane rotor 32 a, facing driveshaft 6. Four branch passages 43 b are formed in vane rotor 32 a in sucha manner as to substantially radially extend from the inner periphery ofcylindrical bore 32 d. Four phase-advance chambers 41 are communicatedwith working-fluid chamber 43 a via respective branch passages 43 b.

On the other hand, the axial passage of second hydraulic line 44 extendsnear the first end face of flow-passage structure 39. Second hydraulicline 44 is further provided with an annular chamber 44 a and a secondworking-fluid passage 44 b. Annular chamber 44 a is formed in the outerperiphery of the cylindrical portion of the first end of flow-passagesstructure 39. Although it is not clearly shown in the drawing, secondworking-fluid passage 44 b has a substantially L shape and formed invane rotor 32 a. Annular chamber 44 a and each of phase-retard chambers42 are communicated with each other via second working-fluid passage 44b.

In the shown embodiment, electromagnetic solenoid-operated directionalcontrol valve 47 is constructed by a four-port, three-position,spring-offset solenoid-actuated directional control valve. Directionalcontrol valve 47 uses a sliding valve spool to change the path of flowthrough the directional control valve. For a given position of the valvespool, a unique flow path configuration exists within the valve.Concretely, directional control valve 47 is designed to switch amongthree positions of the spool, namely a spring-offset position shown inFIG. 6, a block-off position (a center position created due to thebalancing opposing forces, that is, the return spring force and theelectromagnetic force produced by the solenoid), and a fullysolenoid-actuated position. In the spring-offset position, fluidcommunication between first hydraulic line 43 and drain passage 46 isestablished, and fluid communication between second hydraulic line 44and supply passage 45 is established. In the block-off position, fluidcommunication between each of first and second hydraulic lines 43-44 andeach of supply passage 45 and drain passage 46 is blocked. In the fullysolenoid-actuated position, fluid communication between first hydraulicline 43 and supply passage 45 is established, and fluid communicationbetween second hydraulic line 44 and drain passage 46 is established.Switching operation among the three positions of the valve spool ofdirectional control valve 47 is executed responsively to a controlcommand signal generated from the output interface circuitry of ECU 22to the solenoid.

The controller (ECU) 22 is common to both of VEL mechanism 1 and VTCmechanism 2.

Also provided is a lock mechanism (or an interlocking device orinterlocking means) disposed between vane member 32 and housing 34, fordisabling rotary motion of vane member 32 relative to housing 34 bylocking and engaging vane member 32 with housing 34, and for enablingrotary motion of vane member 32 relative to housing 34 by unlocking (ordisengaging) vane member 32 from housing 34. That is, by theinterlocking means, intake valve closure timing IVC of each of intakevalves 4, 4 can be locked or fixed to a predetermined timing value.

As can be seen from the longitudinal cross section of FIG. 6, the lockmechanism (interlocking means) is comprised of a lock-pin sliding-motionpermitting bore (simply, a lock-pin bore) 50, a lock pin 51, anengaging-hole structural member 52 having a substantially C shape inlateral cross section and press-fitted into a through hole formed inrear cover 36, an engaging hole 52 a defined in the C-shapedengaging-hole structural member 52, a spring retainer 53, and a returnspring (a coiled compression spring) 54. Lock-pin bore 50 is formed inthe inverted trapezoidal blade 32 b of the relatively greatercircumferential width (the maximum circumferential width) and formed inrear cover 36, such that lock-pin bore 50 extends in the axial directionof drive shaft 6. Lock pin 51 is slidably accommodated in lock-pin bore50 and has a cylindrical bore closed at one end. A tapered head portion51 a of lock pin 51 is engaged with or disengaged from engaging hole 52a. Spring retainer 53 is fitted into a space defined by the innerperipheral wall surface of front cover 35 and lock-pin bore 51. Returnspring 54 is provided to permanently force lock pin 51 toward theinternal space of engaging hole 52 a. Although it is not clearly shownin FIG. 6, the phase-converter housing structure, constructed by frontand rear covers 35-36 and cylindrical housing 34, is also designed tosupply working oil (hydraulic pressure) in phase-retard chamber 42and/or working oil (hydraulic pressure) discharged from oil pump 49 viaan oil hole formed in the phase-converter housing structure intoengaging hole 52 a.

Lock pin 51 operates to disable relative rotation between timingsprocket 33 and drive shaft 6 by locking and engaging tapered headportion 51 a of lock pin 51 with engaging hole 52 a in a predeterminedposition where vane member 32 reaches its maximum phase-retard position,by way of the spring force of return spring 54. Relative rotationbetween timing sprocket 33 and drive shaft 6 is enabled by unlocking (ordisengaging) tapered head portion 51 a of lock pin 51 from engaging hole52 a by way of the hydraulic pressure delivered from phase-retardchamber 42 and/or oil pump 49 into engaging hole 52 a. That is, taperedhead portion 51 a of lock pin 51 is forced out of engaging hole 52 aunder hydraulic pressure fed into the engaging hole from phase-retardchamber 42 and/or oil pump 49.

As previously described with reference to FIG. 7, two rows of returnsprings 55-56, each of which serves as a biasing device or biasingmeans, are disposed between the spring-retaining-hole equipped backwardsidewall surface of blade 32 b and the spring-retaining sidewall surfaceof shoe 34 a, for permanently biasing the associated blade 32 b (vanemember 32) toward the phase-retard side. In the shown embodiment, returnsprings 55-56 are constructed by coil springs having the same size andthe same spring stiffness.

Although, in FIGS. 7-8, two return springs 55-56 are illustrated to beoverlapped with each other, actually, two return springs 55-56 aredisposed in parallel with each other. As can be seen from the lateralcross section of FIG. 7, the axial length of each of springs 55-56 isdimensioned to be greater than the circumferential distance between thespring-retaining-hole equipped backward sidewall surface of blade 32 band the spring-retaining sidewall surface of shoe 34 a with the blade 32b held at the maximum phase-retard position. Return springs (coilsprings) 55-56 have the same free height.

The distance between the axes of two parallel coil springs 55-56 ispreset to a predetermined distance that the outer peripheries of coilsprings 55-56 are not brought into contact with each other under acondition of maximum compressive deformation of each of coil springs55-56 (see FIG. 8). One end of each of coil springs 55-56, facing theassociated blade 32 b, is retained in a thin-plate spring retainer (notshown) fitted to concave groove (spring retaining hole) 32 c.

Hereinafter described in detail is the basic operation of intake-valveVTC mechanism 2. First, during an engine stopped period, the output ofcontrol current (exciting current) from ECU 22 to the solenoid ofdirectional control valve 47 is stopped. Thus, by means of springs55-56, the valve spool of directional control valve 47 is mechanicallyshifted to its spring-offset position (a default position) at whichfluid communication between first hydraulic line 43 and drain passage 46is established, and simultaneously fluid communication between secondhydraulic line 44 and supply passage 45 is established. Also, when theengine is in its stopped state, the discharge pressure from oil pump 49becomes zero and thus there is no hydraulic pressure supplied to the VTCmechanism 2.

Under these conditions, as shown in FIG. 7, vane member 32 is forcedtoward the maximum phase-retard side by the spring forces of returnsprings 55-56, such that the inverted trapezoidal vane blade 32 b of themaximum circumferential width is brought into abutted-engagement withthe sidewall of shoe 34 a facing phase-advance chamber 41. At the sametime, tapered head portion 51 a of lock pin 51 of the lock mechanism isengaged with the engaging hole 52 a, and thus vane member 32 can bestably held at the maximum phase-retard position. That is, withdirectional control valve 47 kept at its default position (thespring-offset position), intake-valve VTC mechanism 2 can be stably heldat the maximum phase-retard position mechanically by means of the lockmechanism (50, 51, 52 a, 54) and return springs 55-56.

Hereunder described briefly is the operation of intake-valve VTCmechanism 2 under several engine operating conditions. First, during anengine starting period, with the ignition switch turned ON, electricmotor 07 is driven to initiate cranking action for crankshaft 02. At theearly stage of cranking, the output interface circuitry of ECU 22 beginsto generate a control command signal to the solenoid of directionalcontrol valve 47, but the hydraulic pressure of working fluid dischargedfrom oil pump 49 does not yet rise adequately just after the engine hasbeen started. Thus, vane member 32 is still held at the maximumphase-retard position by means of the lock mechanism (50, 51, 52 a, 54)and return springs 55-56.

At this time, the solenoid of directional control valve 47 is held atits default position (the spring-offset position) responsively to acontrol signal from ECU 22, such that fluid communication between secondhydraulic line 44 and supply passage 45 is established and fluidcommunication between first hydraulic line 43 and drain passage 46 isestablished. Under these conditions, on the one hand, owing to a gradualhydraulic pressure rise, hydraulic pressure produced by oil pump 49 issupplied through supply passage 45 and second hydraulic line 44 intoeach of phase-retard chambers 42. On the other hand, there is no supplyof hydraulic pressure to each of phase-advance chambers 41 in the samemanner as the engine stopped state. That is, hydraulic pressure isrelieved from each of phase-advance chambers 41 through first hydraulicline 43 and drain passage 46 into oil pan 48 and thus the hydraulicpressure in each of phase-advance chambers 41 is kept low.

After the hydraulic pressure produced by oil pump 49 has risenadequately, the variable phase control system (intake-valve VTCmechanism 2) enables rapid and accurate vane position control viadirectional control valve 47. In more detail, working fluid, suppliedinto phase-retard chamber 42, is also delivered from phase-retardchamber 42 into engaging hole 52 a, and thus owing to a hydraulicpressure rise of phase-retard chamber 42 a hydraulic pressure inengaging hole 52 a also rises. As a result, lock pin 51 moves backwardsagainst the spring bias of return spring 54 and then tapered headportion 51 a of lock pin 51 is forced out of engaging hole 52 a, so asto permit relative rotation between phase-converter housing 34 and vanemember 32, and thus to enable rapid and accurate vane position control.

For instance, in an engine idling state after the internal combustionengine has been warmed up, the solenoid of directional control valve 47is also held at its default position responsively to a control signalfrom ECU 22, such that fluid communication between second hydraulic line44 and supply passage 45 is established and fluid communication betweenfirst hydraulic line 43 and drain passage 46 is established. Therefore,owing to a rise in hydraulic pressure supplied to each of phase-retardchambers 42, vane member 32 is maintained at the relative position shownin FIG. 7 by the aid of the spring forces of return springs 55-56. Thus,the angular phase of drive shaft 6 relative to timing sprocket 33 isalso maintained at the phase-retard side.

Thereafter, when the engine operating condition is shifted to alow-speed middle-load range, the solenoid of directional control valve47 is shifted to its fully solenoid-actuated position responsively to acontrol signal from ECU 22, such that fluid communication between firsthydraulic line 43 and supply passage 45 is established and fluidcommunication between second hydraulic line 44 and drain passage 46 isestablished.

Hence, hydraulic pressure in each of phase-retard chambers 42 isreturned through second hydraulic line 44 and drain passage 46 into oilpan 48, and thus the hydraulic pressure in each of phase-retard chambers42 becomes low, whereas the hydraulic pressure in each of phase-advancechambers 41 becomes high.

Therefore, owing to a rise in hydraulic pressure supplied to each ofphase-advance chambers 41, vane member 32 is rotated clockwise to therelative position shown in FIG. 8 against the spring forces of returnsprings 55-56. Thus, the angular phase of drive shaft 6 relative totiming sprocket 33 is converted to the phase-advance side. By switchingthe axial position of the spool of directional control valve 47 to theblock-off position (the center position created due to the balancingopposing forces, that is, the return spring force and theelectromagnetic force produced by the solenoid) during rotary motion ofdrive shaft 6 relative to timing sprocket 33, it is possible to hold theangular phase of drive shaft 6 relative to timing sprocket 33 at anarbitrary relative phase angle.

Furthermore, when the engine operating condition is shifted from alow-speed range to a normal middle-speed range, and further shifted to ahigh-speed range, in the same manner as the engine idling state afterengine warm-up, the solenoid of directional control valve 47 iscontrolled to the default position. As a result, a fall in hydraulicpressure in each of phase-advance chambers 41 occurs and simultaneouslya rise in hydraulic pressure supplied to each of phase-retard chambers42 occurs. By a combined force of the supplied hydraulic pressure andthe spring forces of return springs 55-56, the angular phase of driveshaft 6 relative to timing sprocket 33 is converted to the phase-retardside (see FIG. 7).

Prior to detailed explanation of control actions executed by ECU 22, therelationship among crankangle of crankshaft 02, intake-valve open timingIVO, and intake-valve closure timing IVC, obtained by the variable valveactuation device of the first embodiment, at each of #1, #3, #4, and #2engine cylinders, are hereunder described in detail in reference to thecharacteristic diagrams of FIGS. 9A-9D. Hereupon, in the shownembodiment, the firing order of the four engine cylinders of theinternal combustion engine is #1→#3→#4→#2. The characteristic diagramsof FIGS. 9A-9D are based on the assumption that the internal combustionengine is in a stopped state, and thus a central phase angle (e.g., aphase at a peak valve lift) of a valve lift characteristic curve of eachof intake valves 4, 4 is stably kept at the maximum phase-retard side(at the default valve timing) by means of intake-valve VTC mechanism 2.

Now, suppose that control shaft 17 of intake-valve VEL mechanism 1 isrotated to control the lift and working angle characteristic of each ofintake valves 4, 4 to the large intake-valve lift L3 and large workingangle D3 characteristic and also the angular position of crankshaft 02reaches a crankangle near the TDC position on compression stroke at the#1 cylinder. At this time, as a matter of course, each of intake valves4, 4 of the #1 cylinder does not open (see FIG. 9A), but each of intakevalves 4, 4 (controlled to the large intake-valve lift L3 and largeworking angle D3 characteristic) of the #3 cylinder of the next cycleopens (see FIG. 9B) and also each of intake valves 4, 4 of the #4cylinder of the next but one cycle opens (see FIG. 9C). That is, intakevalves 4, 4 of the two different engine cylinders, namely, the #3cylinder and the #4 cylinder, simultaneously open.

Next, check for a specific state where intake valves 4, 4 of all enginecylinders simultaneously close at different degrees of crankshaftrotation. As can be seen from the characteristic diagrams of FIGS.9A-9D, such a specific state does not exist. In other words, acrankangle area, in which (i) a lift and working angle characteristiccurve of each of intake valves 4, 4 of a first cylinder selected fromthe four engine cylinders and (ii) a lift and working anglecharacteristic curve of each of intake valves 4, 4 of a second cylinderselected from the four engine cylinders can be partially overlapped witheach other, exists.

Even as for all crankangles, each of intake valves 4, 4 of at least oneengine cylinder is kept open, and thus a spring reaction force of valvesprings 5, 5 acts on control cam 18. Under these conditions, whenactuating control shaft 17 by motor 20 (the control-shaft actuator)through ball-screw mechanism 21 with the internal combustion engine keptin its stopped state, control shaft 17 cannot be smoothly rotated due tothe spring reaction force acting on control cam 18 and a large staticfriction coefficient of the sliding-contact portion between control cam18 and rocker arm 11.

Next, suppose that control shaft 17 of intake-valve VEL mechanism 1 isrotated to control the lift and working angle characteristic of each ofintake valves 4, 4 to the middle intake-valve lift L2 and middle workingangle D2 characteristic. In a similar manner to a case that theintake-valve lift and working angle characteristic is controlled to thelarge intake-valve lift L3 and large working angle D3 characteristic, acrankangle area, in which (i) an intake-valve lift and working anglecharacteristic curve of a first cylinder selected from the four enginecylinders and (ii) an intake-valve lift and working angle characteristiccurve of a second cylinder selected from the four engine cylinders canbe partially overlapped with each other, exists (see the fine solidlines of the characteristic curves illustrated in FIGS. 9A-9D). Even asfor all crankangles, each of intake valves 4, 4 of at least one enginecylinder is kept open, and thus a spring reaction force of valve springs5, 5 acts on control cam 18. Under these conditions, when actuatingcontrol shaft 17 by motor 20 (the control-shaft actuator) throughball-screw mechanism 21 with the internal combustion engine kept in itsstopped state, control shaft 17 cannot be smoothly rotated due to thespring reaction force acting on control cam 18 and a large staticfriction coefficient of the sliding-contact portion between control cam18 and rocker arm 11, even in the case of the middle working angle D2characteristic, in a similar manner to the large working angle D3characteristic.

Next, suppose that control shaft 17 of intake-valve VEL mechanism 1 isrotated to control the lift and working angle characteristic of each ofintake valves 4, 4 to the small intake-valve lift L1 and small workingangle D1 characteristic. Assuming that the angular position ofcrankshaft 02 reaches a crankangle corresponding to a point “A”indicated by the asterisk in FIGS. 9A-9D, each of intake valves 4, 4 ofthe #1, #4, and #2 cylinders does not open (see FIGS. 9A, 9C, and 9D),but each of intake valves 4, 4 (controlled to the small intake-valvelift L1 and small working angle D1 characteristic) of the #3 cylinderopens (see FIG. 9B). Thus, a spring reaction force of valve springs 5, 5acts on control cam 18. Under these conditions, when actuating controlshaft 17 by motor 20 through ball-screw mechanism 21 with the internalcombustion engine kept in its stopped state, control shaft 17 cannot besmoothly rotated due to the spring reaction force and a large staticfriction coefficient of the sliding-contact portion between control cam18 and rocker arm 11, even in the case that the intake-valve lift andworking angle characteristic is controlled to the small intake-valvelift L1 and small working angle D1 characteristic.

Notice that, in the case that the intake-valve lift and working anglecharacteristic has been controlled to the small intake-valve lift L1 andsmall working angle D1 characteristic, a specific crankangle area (i.e.,a crankangle area α1, a crankangle area α2, a crankangle area α3, and acrankangle area α4), in which (i) a lift and working anglecharacteristic curve of each of intake valves 4, 4 of one cylinderselected from the four engine cylinders and (ii) a lift and workingangle characteristic curve of each of intake valves 4, 4 of the othercylinder selected from the four engine cylinders cannot be overlappedwith each other, exists. Within the specific crankangle area (α1, α2,α3, α4), intake valves 4, 4 of all engine cylinders simultaneouslyclose. Therefore, assuming that the angular position of crankshaft 02reaches a specific crankangle corresponding to a point “B” indicated bythe asterisk in FIGS. 9A-9D and included within the specific crankanglearea α1, there is a less valve-spring reaction force acting on controlcam 18 (or control shaft 17), thus ensuring a smooth rotary motion ofcontrol cam 18. As soon as control cam 18 begins to smoothly rotate, afriction coefficient of the sliding-contact portion between control cam18 and rocker arm 11 changes from a large static friction coefficient toa small kinetic friction coefficient. Hence, in the case that theangular position of crankshaft 02 has been controlled to a crankangleincluded within the specific crankangle area (α1, α2, α3, α4), it ispossible to smoothly change or convert the intake-valve lift and event(working angle) characteristic from the small lift L1 and working angleD1 characteristic to the middle lift L2 and working angle D2characteristic, and further to smoothly change or convert to the largelift L3 and working angle D3 characteristic. Actually, even when theangular position of crankshaft 02 has been adjusted to a specificcrankangle corresponding to the point “B” indicated by the asterisk inFIGS. 9A-9D and included within the specific crankangle area α1,according to the progress of a change (or a conversion) in theintake-valve lift and working angle characteristic to either the middlelift L2 and working angle D2 characteristic or the large lift L3 andworking angle D3 characteristic, of all four engine cylinders, intakevalves 4, 4 of some cylinders (two cylinders in the shown embodiment)begin to open. At this time, by virtue of a transition of the frictioncoefficient of the sliding-contact portion of control cam 18 and rockerarm 11 to a small kinetic friction coefficient and rotational inertia ofcontrol shaft 17, which begins to already rotate, it is possible tomaintain and ensure a good valve lift and event conversionresponsiveness of intake-valve VEL mechanism 1.

That is, the engine control system of the first embodiment isconfigured, so that, during an internal combustion engine stoppingperiod, a crankshaft-rotation stopped position of crankshaft 02 isadjusted to a crankangle included within the specific crankangle area(α1, α2, α3, α4) by means of electric motor 07. At this time, theintake-valve lift and event (working angle) characteristic is controlledto the small lift L1 and working angle D1 characteristic.

On the other hand, during an engine restarting period, a control signalis outputted to motor 20 of intake-valve VEL mechanism 1 for the purposeof converting to a desired working angle by means of the attitudecontrol mechanism (including control shaft 17 and control cam 18) beforeinitiating cranking action of crankshaft 02. Hereby, a valve lift andevent conversion to the desired working angle can be initiated beforecranking the engine. Thus, it is possible to improve a valve lift andevent conversion responsiveness of intake-valve VEL mechanism 1 (byvirtue of a transition of the friction coefficient of thesliding-contact portion between control cam 18 and rocker arm 11 from alarge static friction coefficient to a small kinetic frictioncoefficient), and also to shorten a converting time (a response time)required for conversion to the desired working angle (by virtue ofrotational inertia of control shaft 17 as well as a transition of thefriction coefficient of the sliding-contact portion between control cam18 and rocker arm 11 to a small kinetic friction coefficient).

By the way, a desired working angle, required for each of intake valves4, 4 during an engine starting period, is different depending on acondition on the engine or vehicle, such as an engine temperature. Forinstance, when the engine temperature is very low, intake-valve closuretiming IVC has to approach closer to a timing value near a piston bottomdead center (BDC) position in order to ensure better combustion. In sucha case, a middle working angle D2 is selected as a desired workingangle.

In contrast to the above, when the engine temperature is high, in orderto suppress preignition and starting-period vibrations, a large workingangle (a maximum working angle) D3 is selected as a desired workingangle. When the large working angle D3 is selected, intake-valve closuretiming IVC is greatly phase-retarded with respect to the BDC position,and thus the temporarily drawn-in fresh air is discharged from thecombustion chamber to the intake port, thereby reducing an effectivecompression ratio. Such a decompressing action contributes to asuppression in preignition and/or a reduction in noise and vibrationswhen starting the engine.

During a general engine restarting period that the engine temperature isneither very low nor high, a small working angle (a minimum workingangle) D1 is selected as a desired working angle. When the minimumworking angle D1 is selected, the actual intake-valve lift becomes asmall lift L1 and simultaneously the actual intake-valve working anglebecomes a small working angle D1, and thus it is possible to reduce africtional loss in the valve operating system and thus to ensure asmooth engine speed rise, thereby realizing a good restartability (asmooth and rapid engine restart).

The minimum working angle D1 contributes to a reduced effectivecompression ratio (i.e., a decompressing effect), but leads to adrawback that preignition can be somewhat promoted owing to anintake-air mixing effect, caused by a phase-retarded intake-valve opentiming IVO. For the reasons discussed above, in the case of high enginetemperatures (high engine oil temperatures), the maximum working angleD3 is superior to the minimum working angle D1.

As discussed above, during an engine stopping period, a crankangle ispreset or pre-adjusted within an all-cylinder valve closed period, inother words, an all-cylinder intake-valve-closed crankangle area (α1,α2, α3, α4), in which intake valves 4, 4 of all cylinders are kept intheir non-lifted states, by means of electric motor 07 of the crankposition change mechanism. During an engine restarting period, prior tocranking action of crankshaft 02, a control signal is outputted tointake-valve VEL mechanism 1 so as to achieve a desired working angle(an engine-start desired working angle Dt described later), determinedbased on an engine condition (or an engine/vehicle condition) such asengine temperature (concretely, engine oil temperature or engine coolanttemperature). Therefore, it is possible to shorten a converting timerequired for conversion to the desired working angle.

Details of the control flow executed by ECU 22 are hereunder describedin reference to the flowchart of FIG. 10.

First, at step S1, a check is made to determine, based on the currentengine/vehicle conditions, whether an engine-stop condition issatisfied. In other words, a check is made to determine whether anecessary condition that the ignition switch should be turned OFF issatisfied. In the case of an automatic-engine-stop-restart systemequipped hybrid vehicle, a check is made to determine whether anecessary condition that the engine should be automatically stopped issatisfied. When the answer to step S1 is in the negative (NO), oneexecution cycle terminates without any control action. Conversely whenthe answer to step S1 is in the affirmative (YES), that is, when theengine-stop condition has been satisfied, step S2 occurs.

At step S2, a control signal is outputted from ECU 22 to thecontrol-shaft actuator (motor 20) of intake-valve VEL mechanism 1 insuch a manner as to switch the operating mode of each of intake valvesto the minimum working angle D1 operating mode (i.e., the smallintake-valve lift L1 and working angle D1 characteristic).

At step S3, a control signal is outputted from ECU 22 to electric motor07 included in the crank position change mechanism in such a manner asto control or adjust the angular position of crankshaft 02 to acrankangle included within the previously-discussed all-cylinder valveclosed period (i.e., the all-cylinder intake-valve-closed crankanglearea (α1, α2, α3, α4), for example the crankangle area α1).

At step S4, a check is made to determine whether the actual workingangle of each of intake valves 4, 4 has been controlled to the minimumworking angle D1 by means of intake-valve VEL mechanism 1 and the actualcrankangle of crankshaft 02 has been controlled to a crankangle includedwithin the previously-discussed all-cylinder intake-valve-closedcrankangle area (e.g., α1). When the answer to step S4 is negative (NO),the routine returns from step S4 to step S2. Conversely when the answerto step S4 is affirmative (YES), step S5 occurs.

At step S5, an engine-stop signal is generated from ECU 22.

At step S6 subsequently to step S5, rotation of the engine is actuallystopped.

The internal combustion engine is held in its engine stopped state,until a subsequent engine-restart action is initiated. In the enginestopped state, intake-valve VEL mechanism 1 is held at a defaultposition where the operating mode of each of intake valves 4, 4 isstably kept at its minimum working angle D1 operating mode by the springforce of coil spring 30, and whereby the small intake-valve lift L1 andsmall working angle D1 characteristic can be maintained. In the enginestopped state, intake-valve VTC mechanism 2 is also held at a defaultposition where valve timing (intake-valve open timing IVO andintake-valve closure timing IVC) of each of intake valves 4, 4 is stablykept at the maximum phase-retard position by the spring forces of returnsprings 55-56, and whereby the maximum phase-retard position ofintake-valve VTC mechanism 2 can be maintained. On the other hand,crankshaft 02 is maintained at an angular position corresponding to acrankangle included within the previously-discussed all-cylinderintake-valve-closed crankangle area (e.g., α1), for example a specificcrankangle corresponding to the point “B” indicated by the asterisk inFIGS. 9A-9D and included within the specific crankangle area α1.

At step S7, a check is made to determine whether an engine-restartcondition is satisfied. For instance, in the presence of a requirementof re-acceleration on automatic-engine-stop-restart system equippedhybrid vehicles, ECU 22 determines that such an engine-restart conditionis satisfied. When the answer to step S7 is negative (NO), one executioncycle terminates. Conversely when the answer to step S7 is affirmative(YES), step S8 occurs.

At step S8, the current engine/vehicle conditions, for example, thecurrent value of engine temperature T, detected by the enginetemperature sensor, are read. Thereafter, step S9 occurs.

At step S9, a check is made to determine whether the engine temperatureT (read through step S8) is higher than a predetermined firsttemperature value T1. When the answer to step S9 is negative (NO), thatis, when T≦T1, in other words, when ECU 22 determines that the engine iscold, the routine proceeds to step S10.

At step S10, the middle working angle D2 is set to an engine-startdesired working angle Dt for intake-valve VEL mechanism 1. Thereafter,the routine proceeds to step S14. At the point of time, corresponding tostep S10, valve timing of each of intake valves 4, 4 is stably kept atthe maximum phase-retard position by means of intake-valve IVC mechanism2, and also intake-valve closure timing IVC is controlled to a timingvalue near the BDC position. For instance, intake-valve closure timingIVC of the #1 cylinder exists near the BDC position of the #1 cylinder(in other words, near the TDC position of the #2 cylinder on compressionstroke). Hence, it is possible to set the effective compression ratio toa high value, thus improving combustion during cold-engine operation.

Returning to step S9, when the answer to step S9 is affirmative (YES),that is, when T>T1, the routine proceeds from step S9 to step S11.

At step S11, a check is made to determine whether the engine temperatureT is higher than or equal to a predetermined second temperature valueT2. When the answer to step S11 is affirmative (YES), that is, whenT≧T2, in other words, when ECU 22 determines that the engine temperatureis high, the routine proceeds to step S12.

At step S12, the large working angle (the maximum working angle) D3 isset to engine-start desired working angle Dt for intake-valve VELmechanism 1. At the point of time, corresponding to step S12,intake-valve VTC mechanism 2 is still stably kept at its defaultposition (that is, at the maximum phase-retard position). Thus,intake-valve closure timing IVC of the #1 cylinder is greatlyphase-retarded from the BDC position of the #1 cylinder (in other words,from the TDC position of the #2 cylinder on compression stroke). In asimilar manner, as for the other cylinders, intake-valve closure timingsIVC tend to be greatly phase-retarded from the respective BDC positions.Hence, it is possible to set the effective compression ratio to a lowvalue, thus effectively suppressing the occurrence of preignition byvirtue of such a decompressing action. Owing to high engine oiltemperature (in other words, low-viscosity engine oil) there is apossibility of increased starting-period noise and vibrations, but byvirtue of the decompressing action it is possible to suppress suchstarting-period noise and vibrations.

Conversely when the answer to step S11 is negative (NO), that is, whenT1<T<T2, in other words, when ECU 22 determines that the enginecondition is a general engine-start condition, the routine proceeds fromstep S11 to step S13.

At step S13, the small working angle (the minimum working angle) D1 isset to engine-start desired working angle Dt for intake-valve VELmechanism 1.

After engine-start desired working angle Dt has been determined aspreviously (see steps S10, S12, S13 in FIG. 10), the routine proceeds tostep S14.

At step S14, a control signal is outputted from ECU 22 to thecontrol-shaft actuator (motor 20) of intake-valve VEL mechanism 1 insuch a manner as to switch or convert the operating mode of each ofintake valves to engine-start desired working angle Dt operating modeprior to cranking action of crankshaft 02.

Herein, a specific crankangle indicated by the above-mentioned point “B”in FIGS. 9A-9D and included within the specific crankangle area α1,corresponds to a crankangle at which intake valves 4, 4 of all enginecylinders simultaneously close, and therefore there is a lessvalve-spring reaction force acting on control cam 18 (or control shaft17), thus ensuring a smooth rotary motion of control cam 18. By virtueof the smooth rotary motion of control cam 18, a smooth conversion tothe desired intake-valve lift and event (working angle) characteristiccan be initiated. Furthermore, the sliding-contact portion betweencontrol cam 18 and rocker arm 11 changes from a static friction area (ofa large friction coefficient) to a kinetic friction area (of a smallfriction coefficient). Such a multiplied and combined effect of thesmooth rotary motion of control cam 18 (the smooth conversion to thedesired intake-valve lift and event (working angle) characteristic) andthe friction-coefficient change of the sliding-contact portion from alarge static friction coefficient to a small kinetic frictioncoefficient, ensures a more smooth lift-and-event (working angle)characteristic conversion of intake-valve VEL mechanism 1. Even during aconversion from the minimum lift L1 and working angle D1 characteristicto the middle lift L2 and working angle D2 characteristic during whichintake valves 4, 4 of some cylinders (two cylinders in the shownembodiment) begin to open or even during a further conversion to themaximum lift L3 and working angle D3 characteristic, thepreviously-discussed smooth lift-and-event (working angle)characteristic conversion action of intake-valve VEL mechanism 1 can becontinuously executed. Moreover, by virtue of rotational inertia ofcontrol shaft 17, which begins to already rotate, the smooth conversionaction can be continuously executed.

At step S15, cranking action of crankshaft 02 is initiated by means ofelectric motor 07. The timing of initiation of cranking action may beset to a point of time when engine-start desired working angle Dt forintake-valve VEL mechanism 1 has been reached. Alternatively, thecranking action may be initiated during a time period from the starttime of conversion to engine-start desired working angle Dt to thefinish time of conversion to engine-start desired working angle Dt, inother words, under an unconfirmed state of completion of conversion toengine-start desired working angle Dt.

In the case of the former, that is, in the confirmed state wherecompletion of conversion to engine-start desired working angle Dt hasbeen confirmed, as a matter of course, at the early stage of crankingaction the lift and event (working angle) characteristic has alreadybeen controlled or adjusted to the engine-start desired working angleDt. Thus, it is possible to provide a desirable high startability(restartability) of the internal combustion engine. Additionally, at thepoint of time when cranking action is initiated by electric motor 07, ithas been already past the time of the peak current of the control-shaftactuator (motor 20) of VEL mechanism 1. Thus, the adequate batteryvoltage can be supplied to electric motor 07, thereby enablingsatisfactory cranking action with much electric energy.

Conversely in the case of the latter, that is, in the unconfirmed stateof completion of conversion to engine-start desired working angle Dt,cranking action can be quickly initiated before completion of conversionto engine-start desired working angle Dt. That is, immediately when stepS7 of FIG. 10 determines that the engine-restart condition is satisfied,cranking action can be quickly initiated, thus enabling a rapidtransition to a burning state of the engine. This means a merit of arapid acceleration rate of the vehicle, for instance, when acceleratingfrom vehicle standstill. The flowchart illustrated in FIG. 10, is mainlybased on the assumption that cranking action can be quickly initiatedunder the unconfirmed state of completion of conversion to engine-startdesired working angle Dt. Hence, through step S15, cranking action isquickly initiated in the unconfirmed state of completion of conversionto engine-start desired working angle Dt.

When taking account of the initial compression of the early stage ofcranking action, the initial compression in the #1 cylinder takes placeat a specific crankangle indicated by the point “B” in FIGS. 9A-9D (thatis, slightly before the TDC position on compression stroke). Therefore,atmospheric pressure has already been flown via the piston-to-cylinderclearance space into the engine cylinder after the engine has beenstopped, and hence the air-fuel mixture is compressed under atmosphericpressure as an initial condition, and then the compression increasesduring a period from the specific crankangle indicated by the point “B”in FIGS. 9A-9D to the TDC position on compression stroke.

However, in the first place, the specific crankangle indicated by thepoint “B” is near the TDC position on compression stroke (exactly,slightly before the TDC position on compression stroke), and thus apiston stroke to the TDC position is very short. This means a slightwork of compression, in other words, a smooth cranking speed increase.From this viewpoint, it is possible to more greatly enhance the enginestartability (the engine restartability). Suppose that the specificcrankangle indicated by the point “B” is just after the TDC position oncompression stroke. The initial compression in the #1 cylinder itselfdoes not take place. In such a case, it is possible to more smoothlyincrease the cranking speed.

Hence, as a matter of course, preignition and starting-periodnoise/vibrations can be effectively suppressed, and also speedy startingaction can be realized.

At step S16 subsequently to step S15, a check is made to determinewhether engine-start desired working angle Dt for intake-valve VELmechanism 1 has been reached. When the answer to step S16 is affirmative(YES), the routine proceeds to step S17.

At step S17, complete explosion control (fuel injection and ignition) isexecuted, and thus certain and speedy starting action has beencompleted.

Returning to step S16, when the answer to step S16 is negative (NO),that is, when engine-start desired working angle Dt is not yet reached,the routine returns from step S16 to step S14, to generate a controlsignal corresponding to engine-start desired working angle Dt once againthrough step S14, and also to continually execute cranking actionthrough step S15. Such a flow returning from step S16 back to step S14is repeatedly executed until engine-start desired working angle Dt hasbeen reached.

By the way, in the first embodiment, as an engine condition (or anengine/vehicle condition), engine temperature (concretely, engine oiltemperature and/or engine coolant temperature) is used. In addition tothe engine temperature, a vehicle speed may be added to theengine/vehicle condition. By virtue of the use of more informationaldata about the engine/vehicle condition, containing vehicle speed aswell as engine temperature, it is possible to more accurately set ordetermine a desired working angle (i.e., engine-start desired workingangle Dt).

Second Embodiment

Referring now to FIGS. 11A-11B and 12, there is shown the engine controldevice of the second embodiment wherein the inventive concept is appliedto an in-line two-cylinder internal combustion engine. The basicconfigurations of intake-valve VEL mechanism 1 and intake-valve VTCmechanism 2, both constructing the variable valve actuation deviceincorporated in the engine control system of the second embodiment, arethe same as those of the first embodiment.

As can be seen from the characteristic diagrams of FIGS. 11A-11B, in theengine stopped state, intake-valve VEL mechanism 1 is held at a defaultposition where the operating mode of each of intake valves 4, 4 is keptat its minimum working angle D1′ operating mode.

As clearly seen in FIGS. 11A-11B, there is a specific crankangle areaα1′, in which intake valves 4, 4 of the two engine cylinders (#1 and #2cylinders) are kept simultaneously in their closed states, near the TDCposition of the #1 cylinder on compression stroke. The specificcrankangle area α1′ corresponds to an interval ranging from a crankangleat which intake valves 4, 4 of the #1 cylinder close (i.e., intake-valveclosure timing IVC of the #1 cylinder) to a crankangle at which intakevalves 4, 4 of the #2 cylinder open (i.e., intake-valve open timing IVOof the #2 cylinder). The length of the specific crankangle area α1′ (seeFIGS. 11A-11B) of the second embodiment, wherein the inventive conceptis applied to a two-cylinder engine, is sufficiently enlarged ascompared to the length of the specific crankangle area α1 (see FIGS.9A-9D) of the first embodiment, wherein the inventive concept is appliedto a four-cylinder engine. This is because the interval between twoengine cylinders in a four-cylinder engine is 180 degrees of crankangle,whereas the interval between two engine cylinders in a two-cylinderengine is 360 degrees of crankangle.

Therefore, in the second embodiment, the control area of a desiredcrankangle to be achieved by the crank position change mechanism can beenlarged to the specific crankangle area α1′. Even when the crankposition change mechanism has a somewhat low control accuracy, thecontrol system of the second embodiment can easily achieve the desiredcrankangle, because of the enlarged specific crankangle area α1′.

Additionally, in the second embodiment, the minimum working angle D1′(that is, a valve open period from intake-valve open timing IVO tointake-valve closure timing IVC at the minimum working angle operatingmode) is assumed to be shorter than the minimum working angle D1 of thefirst embodiment, that is, D1′<D1.

The minimum working angle D1 of the first embodiment, obtained byintake-valve VEL mechanism 1, means an effective working angle (aneffective lifted period) from an actual valve-open point to an actualvalve-closed point, fully taking account of the valve clearance Δ (seeFIG. 5). On the other hand, as shown in FIG. 12, the minimum workingangle D1′ of the second embodiment, obtained by intake-valve VELmechanism 1, means a really effective lift section ranging from thepoint immediately after the leading edge of the positive valve openingacceleration (except a moderate valve-opening ramp section (i.e., a verysmall lift ΔL), which permits moderate valve movement in the first stage(the initial stage) of opening motion of intake valve 4) to the pointimmediately before the trailing edge of the positive valve closingacceleration (except a moderate valve-closing ramp section (i.e., a verysmall lift ΔL), which permits moderate valve movement in the last stageof closing motion of intake valve 4).

As can be appreciated from the relationship defined by the inequality ofD1′<D1, the length of the specific crankangle area α1′ (see FIGS.11A-11B) of the second embodiment can be further enlarged in comparisonwith the length of the specific crankangle area α1 (see FIGS. 9A-9D) ofthe first embodiment, by the difference (D1−D1′) between the minimumworking angle D1 and the minimum working angle D1′. As a result of this,the level of the required control accuracy of the crank position changemechanism can be relaxed or reduced. In other words, owing to theenlarged control area of a desired crankshaft-rotation stopped position,the controllability of the crank position change mechanism can begreatly improved.

FIG. 12 shows the linkage attitude of the multinodular-link motiontransmitting mechanism (the motion converter) of intake-valve VELmechanism 1, during a ramp-lift period. As seen from the axial view ofFIG. 12, during the ramp-lift period, the offset distance of the pointof action of force (a load Fs) from the center X of oscillating motionof rockable cam 9 becomes a sufficiently small offset distance ΔT. Evenwhen load Fs (the spring force of valve spring 5) acts on rockable cam9, a moment acting on rockable cam 9 also becomes a sufficiently smallmoment ΔM, because of the sufficiently small offset distance ΔT. Hence,the magnitude of load transmitted via link rod 13 and acting on controlcam 18 also becomes sufficiently small. This enables a smooth rotarymotion of control shaft 17 even during the engine stopping period. Thatis, it is possible to realize a smooth operating mode shift to a desiredintake-valve lift and event (working angle) characteristic under asubstantially zero-lift state.

For instance, in the case of a small lift L1 exceeding the ramp liftperiod, as shown in FIG. 3B, the offset distance of the point of actionof load Fs from the center X of oscillating motion of rockable cam 9becomes a large offset distance T. A moment acting on rockable cam 9also becomes a large moment M, because of the large offset distance T.Hence, the magnitude of load transmitted via link rod 13 and acting oncontrol cam 18 becomes remarkably large, and thus it is difficult tosmoothly rotate control shaft 17 under an engine stopped state.

Furthermore, suppose that a midpoint “B′” (indicated by the asterisk inFIGS. 11A-11B) of the control area of a desired crankangle to beachieved by the crank position change mechanism (that is, the specificcrankangle area α1′) is set as a target point (a desiredcrankshaft-rotation stopped position of crankshaft 02), and thencrankshaft rotation position control (simply, crank position control) isperformed by electric motor 07 of the crank position change mechanism.In this case, regardless of each piston's individual operatingcharacteristics, for example a difference in frictional resistancesbetween reciprocating pistons 01, it is possible to accurately controlor adjust the angular position of crankshaft 02 to a crankangle includedwithin the desired-crankangle control area (i.e., the specificcrankangle area α1′).

As will be appreciated from the above, the inventive concept of theinvention is applied to a four-cycle four-cylinder internal combustionengine (the first embodiment) and also to a two-cylinder internalcombustion engine (the second embodiment). It will be understood thatthe invention is not limited to the particular embodiments shown anddescribed herein, but that various changes and modifications may bemade. For instance, the number of engine cylinders is not limited to “4”or “2”. As appreciated, the all-cylinder intake-valve-closed crankanglearea (e.g., the specific crankangle area α1 (or α1′) in which intakevalves 4, 4 of all cylinders are kept in their valve-closed states)tends to be enlarged, as the number of engine cylinders reduces (forexample, 6-cylinder engine→5-cylinder engine→4-cylinderengine→3-cylinder engine→2-cylinder engine). Thus, the fewer the numberof engine cylinders, the easier the crank position control is achieved.

In the shown embodiments, the crank position change mechanism isconstructed mainly by electric motor 07 and pinion gear mechanism 06. Inlieu thereof, an electric motor may be directly connected to the rearend of crankshaft 02.

As a vehicle to which the engine control device (or the engine controlsystem) of the embodiment is applied, an idling-stop system andmulti-cylinder internal-combustion-engine equipped vehicle isexemplified. In this case, the vehicle also stops simultaneously with anautomatic stop of the engine. The inventive concept can be applied to anautomatic engine stop-restart system equipped hybrid vehicle, which canbe propelled by means of a motor/generator or an electric motor (servingas a propelling power source) in an engine stopped state.

In the shown embodiments, the initial attitude of rockable cam 9 ischanged via the multinodular-link motion transmitting mechanism (themotion converter) of intake-valve VEL mechanism 1, constructing part ofthe variable valve actuation device, by rotary motion of control shaft17. Instead of using such a rotary motion of control shaft 17, themotion converter may be configured such that the initial attitude ofrockable cam 9 is changed by displacing an axial position of controlshaft 17.

OPERATIONS AND EFFECTS

(1) According to the internal combustion engine control device of theshown embodiments, when the engine stops, the variable valve actuationdevice is controlled to change an operating mode of each of the intakevalves to a specific state where an all-cylinder valve closed period,during which the intake valves of the cylinders are all kept in theirnon-lifted states, occurs, and the crank position change mechanism iscontrolled to change the crankshaft-rotation stopped position to acrankangle included within the all-cylinder valve closed period. Whenrestarting the engine, the variable valve actuation device is controlledto bring the operating mode of each of the intake valves closer to anengine-start desired lift characteristic Dt suited to an enginecondition (e.g., engine temperature T), prior to cranking action. Thisensures a good engine restartability.

(2) Concretely, in the shown embodiments, when the engine stops, thevariable valve actuation device (in particular, intake-valve VELmechanism 1) is controlled to change the working angle of the intakevalve to a small working angle side within a controllable working anglerange.

(3) Also, the variable valve actuation device (in particular,intake-valve VEL mechanism 1) is configured so that a lower limit of acontrollable working angle range of the intake valve can be controlledto a minimum working angle D1 that does not reach a zero working angle.

Thus, conversion into a desired working angle (i.e., engine-startdesired working angle Dt) starts from an initial working angle (e.g.,the minimum working angle D1), which does not yet reach a zero workingangle. Owing to such a narrow width of conversion, it is possible toshorten a response time of conversion into the desired working angle,and also to enables a more simplified configuration of the variablevalve actuation device.

(4) As may be appreciated from the flowchart of FIG. 10 and thecharacteristic diagrams of FIGS. 9A-9D, when engine temperature T,detected when restarting the engine, is lower than or equal to apredetermined first temperature value T1, valve closure timing IVC ofthe intake valve is controlled to a timing value near a piston bottomdead center position on intake stroke. Thus, in the case that enginetemperature T, detected when restarting the engine, is very low (T≦T1),it is possible to improve combustion.

(5) Additionally, when the engine temperature T, detected whenrestarting the engine, is higher than or equal to a predetermined secondtemperature value T2 exceeding the first temperature value T1, the valveclosure timing IVC of the intake valve is controlled to a timing valuephase-retarded from the piston bottom dead center position on intakestroke. Thus, in the case that engine temperature T, detected whenrestarting the engine, is high (T2≦T), it is possible to suppresspreignition and starting-period noise/vibrations.

(6) Furthermore, an intake-valve working angle (e.g., D2), suited to therestarting-period engine temperature lower than or equal to the firsttemperature value T1, is controlled to be greater than an intake-valveworking angle (e.g., D1), suited to the restarting-period enginetemperature higher than the first temperature value T1 and lower thanthe second temperature value T2, and also controlled to be less than anintake-valve working angle (e.g., D3), suited to the restarting-periodengine temperature higher than the second temperature value T2.

(7) When the engine stops, the crankshaft-rotation stopped position ischanged to a crankangle included within the all-cylinder valve closedperiod by the crank position change mechanism, after the variable valveactuation device has been controlled to change the operating mode ofeach of the intake valves to the specific state where the all-cylindervalve closed period, during which the intake valves of the cylinders areall kept in their non-lifted states, occurs.

(8) The internal combustion engine may be mounted on an automotivevehicle having an idling-stop function.

(9) The crank position change mechanism may be configured to change thecrankshaft-rotation stopped position by controlling an electric motorused as a propelling power source of the vehicle.

(10) The internal combustion engine may be mounted on a hybrid vehiclethat can be propelled by only an electric motor, under a stopped stateof the engine.

(11) The crank position change mechanism may be configured to change thecrankshaft-rotation stopped position by controlling an alternator thatconverts mechanical energy into electrical energy for charging abattery.

(12) When restarting the engine, the variable valve actuation device maybe controlled to bring the operating mode of each of the intake valvescloser to the engine-start desired lift characteristic (i.e.,engine-start desired working angle Dt) suited to the engine condition(e.g., engine temperature T), immediately after an electric power sourcehas been switched ON with an ignition switch turned ON.

(13) When initiating cranking action forcibly by means of the ignitionswitch before controlling the variable valve actuation device to bringthe operating mode of each of the intake valves closer to theengine-start desired lift characteristic (i.e., engine-start desiredworking angle Dt) suited to the engine condition (e.g., enginetemperature T), the variable valve actuation device may be controlled tobring the engine-start desired lift characteristic of the intake valveduring the forcible cranking action.

(14) Alternatively, The cranking action may be initiated, after theengine-start desired lift characteristic (i.e., engine-start desiredworking angle Dt) of the intake valve has been reached via the variablevalve actuation device.

(15) The control shaft 17 may be configured to be driven directly by adriving force produced by an electric motor.

(16) A specific crankangle area, including a specific crankangle (e.g.,point “B”) corresponding to the crankshaft-rotation stopped position tobe reached when the engine stops, may be set to a given crankangle area(e.g., α1) including a piston top dead center position on compressionstroke with respect to a given engine cylinder (e.g., #1 cylinder).

(17) In lieu thereof, a specific crankangle area, including a specificcrankangle corresponding to the crankshaft-rotation stopped position tobe reached when the engine stops, may be set to a given crankangle area(e.g., α1) including a crankangle just before a piston top dead centerposition on compression stroke with respect to a given engine cylinder(e.g., #1 cylinder).

In the case of the engine control device (the engine control system) asdescribed in the above items (16)-(17), during a period from the timewhen the engine stops to the time when the engine is restarted, anin-cylinder pressure in the given engine cylinder (e.g., #1 cylinder) inclose proximity to or just before a piston top dead center position oncompression stroke, tends to lower to atmospheric pressure, but adecompressing effect can be realized owing to a less piston stroke or avery short piston stroke during a first compressing action. Therefore,it is possible to more remarkably suppress or reduce preignition andstarting-period noise and vibrations.

The entire contents of Japanese Patent Application No. 2009-238810(filed Oct. 16, 2009) are incorporated herein by reference.

While the foregoing is a description of the preferred embodimentscarried out the invention, it will be understood that the invention isnot limited to the particular embodiments shown and described herein,but that various changes and modifications may be made without departingfrom the scope or spirit of this invention as defined by the followingclaims.

1. An internal combustion engine control device comprising: a crankposition change mechanism configured to change a crankshaft-rotationstopped position of a crankshaft of an internal combustion engine; and avariable valve actuation device configured to change at least a workingangle of each of intake valves of a plurality of engine cylinders bychanging a position of a control shaft, wherein, when the engine stops,the variable valve actuation device is controlled to change an operatingmode of each of the intake valves to a specific state where anall-cylinder valve closed period, during which the intake valves of thecylinders are all kept in their non-lifted states, occurs, and the crankposition change mechanism is controlled to change thecrankshaft-rotation stopped position to a crankangle included within theall-cylinder valve closed period, and wherein, when restarting theengine, the variable valve actuation device is controlled to bring theoperating mode of each of the intake valves closer to an engine-startdesired lift characteristic suited to an engine condition, prior tocranking action.
 2. The internal combustion engine control device asclaimed in claim 1, wherein: when the engine stops, the variable valveactuation device is controlled to change the working angle of the intakevalve to a small working angle side within a controllable working anglerange.
 3. The internal combustion engine control device as claimed inclaim 1, wherein: when an engine temperature, detected when restartingthe engine, is lower than or equal to a predetermined first temperaturevalue, valve closure timing of the intake valve is controlled to atiming value near a piston bottom dead center position on intake stroke.4. The internal combustion engine control device as claimed in claim 3,wherein: when the engine temperature, detected when restarting theengine, is higher than or equal to a predetermined second temperaturevalue exceeding the first temperature value, the valve closure timing ofthe intake valve is controlled to a timing value phase-retarded from thepiston bottom dead center position on intake stroke.
 5. The internalcombustion engine control device as claimed in claim 4, wherein: anintake-valve working angle, suited to the restarting-period enginetemperature lower than or equal to the first temperature value, iscontrolled to be greater than an intake-valve working angle, suited tothe restarting-period engine temperature higher than the firsttemperature value and lower than the second temperature value, and alsocontrolled to be less than an intake-valve working angle, suited to therestarting-period engine temperature higher than the second temperaturevalue.
 6. The internal combustion engine control device as claimed inclaim 1, wherein: when the engine stops, the crankshaft-rotation stoppedposition is changed to a crankangle included within the all-cylindervalve closed period by the crank position change mechanism, after thevariable valve actuation device has been controlled to change theoperating mode of each of the intake valves to the specific state wherethe all-cylinder valve closed period, during which the intake valves ofthe cylinders are all kept in their non-lifted states, occurs.
 7. Theinternal combustion engine control device as claimed in claim 1,wherein: the internal combustion engine is mounted on an automotivevehicle having an idling-stop function.
 8. The internal combustionengine control device as claimed in claim 7, wherein: the crank positionchange mechanism is configured to change the crankshaft-rotation stoppedposition by controlling an electric motor used as a propelling powersource of the vehicle.
 9. The internal combustion engine control deviceas claimed in claim 1, wherein: the crank position change mechanism isconfigured to change the crankshaft-rotation stopped position bycontrolling an electric motor.
 10. The internal combustion enginecontrol device as claimed in claim 1, wherein: the internal combustionengine is mounted on a hybrid vehicle that can be propelled by only anelectric motor, under a stopped state of the engine.
 11. The internalcombustion engine control device as claimed in claim 1, wherein: thecrank position change mechanism is configured to change thecrankshaft-rotation stopped position by controlling an alternator thatconverts mechanical energy into electrical energy for charging abattery.
 12. The internal combustion engine control device as claimed inclaim 1, wherein: when restarting the engine, the variable valveactuation device is controlled to bring the operating mode of each ofthe intake valves closer to the engine-start desired lift characteristicsuited to the engine condition, immediately after an electric powersource has been switched ON with an ignition switch turned ON.
 13. Theinternal combustion engine control device as claimed in claim 12,wherein: when initiating cranking action forcibly by means of theignition switch before controlling the variable valve actuation deviceto bring the operating mode of each of the intake valves closer to theengine-start desired lift characteristic suited to the engine condition,the variable valve actuation device is controlled to bring theengine-start desired lift characteristic of the intake valve during theforcible cranking action.
 14. The internal combustion engine controldevice as claimed in claim 12, wherein: the cranking action isinitiated, after the engine-start desired lift characteristic of theintake valve has been reached via the variable valve actuation device.15. The internal combustion engine control device as claimed in claim14, wherein: the control shaft is driven directly by a driving forceproduced by an electric motor.
 16. The internal combustion enginecontrol device as claimed in claim 1, wherein: a specific crankanglearea, including a specific crankangle corresponding to thecrankshaft-rotation stopped position to be reached when the enginestops, is set to a given crankangle area including a piston top deadcenter position on compression stroke with respect to a given enginecylinder.
 17. The internal combustion engine control device as claimedin claim 1, wherein: a specific crankangle area, including a specificcrankangle corresponding to the crankshaft-rotation stopped position tobe reached when the engine stops, is set to a given crankangle areaincluding a crankangle just before a piston top dead center position oncompression stroke with respect to a given engine cylinder.
 18. Aninternal combustion engine control device comprising: a crank positionchange mechanism configured to change a crankshaft-rotation stoppedposition of a crankshaft of an internal combustion engine; and avariable valve actuation device configured to change at least a workingangle of each of intake valves of a plurality of engine cylinders bychanging a position of a control shaft, wherein, when the engine stops,the crank position change mechanism, together with the variable valveactuation device, is controlled to execute crank position control aswell as intake-valve operating characteristic control in such a manneras to realize a specific state where there is a less valve-springreaction force acting on the control shaft, and wherein, when restartingthe engine, the variable valve actuation device is controlled to bringthe position of the control shaft closer to a desired position suited tostart-up of the engine, prior to cranking action.
 19. The internalcombustion engine control device as claimed in claim 18, wherein: thevariable valve actuation device is configured so that a lower limit of acontrollable working angle range of the intake valve can be controlledto a minimum working angle that does not reach a zero working angle. 20.An internal combustion engine control system comprising: a crankposition change mechanism configured to change a crankshaft-rotationstopped position of a crankshaft of an internal combustion engine; avariable valve actuation device configured to change a valve lift aswell as a working angle of each of intake valves of a plurality ofengine cylinders; and a controller configured to control operations ofthe crank position change mechanism and the variable valve actuationdevice, wherein, when the engine stops, the controller controls thevariable valve actuation device to change an operating characteristic ofeach of the intake valves to a specific state where an all-cylindervalve closed period, during which the intake valves of the cylinders areall kept in their non-lifted states, occurs, and the controller controlsthe crank position change mechanism to change the crankshaft-rotationstopped position to a crankangle included within the all-cylinder valveclosed period, and wherein, when restarting the engine, the controllercontrols the variable valve actuation device to bring an engine-startdesired lift characteristic of each of the intake valves, suited to anengine condition, prior to cranking action.